Kimpex Arrow Skis Installation Definition

Publication number US3 A1 Publication type Application Application number US 11/623,873 Publication date Aug 30, 2007 Filing date Jan 17, 2007 Priority date Feb 24, 2006 Also published as Publication number 11623873, 623873, US 20 A1, US 2007/199753 A1, US 3 A1, US 3A1, US A1, US A1, US-A1-3, US-A1-, US20A1, US2007/199753A1, US3 A1, US3A1, US A1, USA1 Inventors, Original Assignee Export Citation,, (22), (22), (5), (2) External Links:. A snowmobile rear suspension is shown comprised of a linear force element (LFE) positioned outside the envelope of the snowmobile endless track. The snowmobile frame includes a chassis mounted to a tunnel, where the chassis includes a central mounting bracket. The LFE is attached at one end to the bracket and at the other end to a bell crank. The bell crank is operatively connected to the slide rails. When the slide rails collapse in normal operation, the bell crank strokes the LFE, and the suspension is progressive throughout the range.

The rear suspension of a snowmobile supports an endless track driven by the snowmobile engine to propel the machine. The track is supported beneath a vehicle chassis by a suspension that is designed to provide a comfortable ride and to help absorb the shock of the snowmobile crossing uneven terrain.

2017-11-07 2017-11-07 2017-11-07. Pit Area Defined. Required Equipment. Authorized Personnel in the Pits. Number of Crew Members in the Hot Pit. To have roll cages; however, members are encouraged to install roll. Pilot Bearing – A bearing, generally within one end of a shaft, which.

Most modern snowmobiles use a slide rail suspension which incorporates a pair of slide rails along with several idler wheels to support the track in its configuration. The slide rails are typically suspended beneath the chassis by a pair of suspension arms, with each arm being attached at its upper end to the chassis of the snowmobile, and at its lower end to the slide rails. The mechanical linkage of the slide rails to the suspension arms and to the snowmobile chassis typically is provided by springs and at least one element acting along a linear path, such as a shock absorber, damper, air shock, shock and spring combination, or other linear force element (LFE).

The springs are loaded to bias the slide rails downwardly away from the snowmobile chassis and the shock absorbers; dampers or LFEs provide damping forces for ride comfort. There are presently two general types of snowmobile rear suspensions in all of the snowmobile industry: coupled and uncoupled. The term “coupled” is given to suspensions that have dependant kinematics front-to-rear and/or rear-to-front (relative to the rear suspension rail beam). That is, a suspension is coupled “front-to-rear” when the front of the suspension is deflected vertically and the rear also moves vertically to some degree. A suspension is coupled “rear-to-front” when the rear of the suspension is deflected vertically and the front also moves vertically to some degree.

A suspension is considered to be coupled “tighter” front-to-rear, or increased coupling bias to the front, if a front deflection causes near the same rear deflection. The same is true if a suspension is coupled “tighter” rear-to-front, or increased coupling bias to the rear: a rear deflection causes near the same front deflection. An uncoupled rear suspension functions independently front-to-rear and rear-to-front. A deflection of the front portion of the suspension causes little to no deflection of the rear portion and vice versa. There are two main advantages to a coupled suspension. First, a coupled suspension shares rate when coupled. There is a distinct rate associated with the front of the suspension and a separate distinct rate associated with the rear of the suspension; when a suspension “couples” it borrows the rate of both the front and rear of the suspension so the overall rate becomes higher than could have been achieved without coupling.

Second, coupling is used to control weight transfer during acceleration. An uncoupled suspension will allow excessive chassis pitch due to the independence of the suspension. Coupling stops this by limiting the angle of the slide rail and by increasing the rate of the suspension and “locking” the suspension geometry. There are many ways to create a coupled rear suspension. The simplest form of a rear suspension is a four-link suspension created by the chassis, two arms, and the slide rails all connected with rotational pivots.

This type of suspension yields only one degree of freedom. The slide rail motion and suspension kinematics are predefined by the length of the 4 links and cannot be altered due to the location of the input (front, rear, or between). This is not desirable because the slide rail will not follow undulating terrain or allow any angle change relative to the chassis due to acceleration. To fix this problem with a basic four-link, one of the links is allowed to change length to some degree. The geometry of the four-link therefore changes relative to the location of the input. A deflection of the front portion of the suspension yields one distinct four-link geometry and a deflection of the rear portion of the suspension yields different distinct four-link geometry. There is always some degree of uncoupled behavior in a coupled suspension when the geometry is not locked front-to-rear or rear-to-front.

It is important to note that most coupling is focused on rear-to-front to help control excessive weight transfer. The majority of differences in rear suspension architecture are driven by accomplishing this same goal of a “sloppy” four-link in different ways. 1 illustrates an example of a traditional rear suspension 10 (illustratively a 2D model of the Polaris Fusion® snowmobile rear suspension design) having slide rails 12, a front suspension arm 14 and a rear suspension arm 16. Front suspension arm 14 is coupled to the slide rails 12 by pivot connection 18.

An opposite end of front suspension arm 14 is pivotably coupled to the chassis. Rear suspension arm 16 is pivotably coupled to the slide rails 12 by pivot connection 20. An opposite end of the rear suspension arm 16 is pivotably coupled to the chassis. Torsion springs are illustratively mounted between the rear torque arm and slide rails 12. First and second linear force elements (LFE) 22 and 24 are coupled between the first and second suspension arms 14 and 16, respectively and the slide rails 12 in a conventional manner.

The coupling bias behavior as described above is dependant on this four-link geometry and is important to rear suspension rate, impact harshness, and ride quality. For example, a perfectly symmetric four-link (A=B and C=D, A parallel to B and C parallel to D) will yield a rail angle that is maintained at the same angle throughout travel. In other words, the rail 12 does not rotate relative to the chassis as the suspension is compressed. This type of movement is not desirable due to the need to achieve traction on undulating terrain.

As deviations to this symmetric geometry are made, the rail angle will change throughout suspension travel. Simply moving the rear point of a conventional suspension upward is not feasible.

The rear arm needs to become significantly shorter than the front. Typical link ratios (A/B) on conventional suspensions are between 1 to 1.5. Ratios other than this are not feasible or do not package in current design envelopes. However, to accommodate a higher rear mount, A/B ratios need to increase to the range of about 1.6 to 2.0. Therefore, in an illustrated embodiment, A/B ratios are preferably 1.6 to 2.0 or greater in coupled suspensions.

6 illustrates the difference in a rear load case coupling angle between a conventional suspension labeled as “Prior Art” (illustratively the Polaris IQ 440 suspension) and the illustrated embodiment described below (labeled as “Improved Rear Suspension Coupled” and “Improved Rear Suspension Uncoupled”). 7 illustrates the difference in the coupling angle between a conventional suspension and the suspension of the present improved suspension invention described below. Conventional suspensions yield a front coupling angle that increases through travel.

This means that as the conventional suspension is compressed, the angle of the slide rail increases. This type of behavior is not ideal because as the rail angle increases, rate and damper velocities decrease ultimately resulting in a regressive suspension.

More desirable is a rail angle that decreases as the suspension is compressed; thus, effectively making the suspension rate progressive (the more regressive the rail angle, the more progressive the rate). However, an increasing coupling angle is difficult to eliminate due to the packaging of a traditional snowmobile suspension. In the illustrated embodiment of the present invention, unconventional packaging of the suspension components results in a vertical difference between the front arm and rear arm chassis mounts of preferably 20% or more of the chassis link length (D) which results in a decreasing coupling angle. Further examination of coupling behavior yields two constraints necessary to maintain reasonable component loads and basic function of the rail/ground interface.

Como Hablar En Publico Sin Temor Pdf Free on this page. First, this angle should be positive. In other words, when a load is applied to the front of the suspension as illustrated by arrow 25 in FIG.

3, the front portion of the slide rails 12 moves more than the rear portion and vice versa for a load applied to the rear of the suspension. Second, there should be no inflections, or change in sign of the slope, in the curve of rail angle versus vertical deflection, as shown in FIGS. In other words, when a load is applied to the front of the suspension, at no point should the rear of the suspension begin to move faster than the front and vice versa for a load applied to the rear of the suspension. Because an uncoupled suspension does not form a distinct four-link, no over-centering can occur.

No link ratio is then necessary for a rear load case in an uncoupled suspension. This is very beneficial, but excessive vehicle pitch and lack of vertical rate usually make uncoupled suspensions behave poorly for load carrying capacity and ride quality. Typically, for these suspensions a link ratio is then tuned only for the front load case. The shock/spring ratio can be tuned to help counteract the deficiencies of an uncoupled suspension. In this way, the rear arm geometry is tuned exclusively to maximize rear load case rate.

Therefore, linkage arm length ratios are tuned for front coupling and rear rate in uncoupled suspensions. As discussed above, the majority of snowmobile rear suspension architectures utilize a combination of springs, dampers, or other similar linear force elements (LFE), all packaged within the envelope of the track. Regardless of how these elements are packaged, these designs typically use two methods to generate vertical rate: 1) the LFE is located so that there is some vertical component reacted between the suspension arm and rail beam, and 2) the LFE is connected to the suspension arm such that a torque reaction is generated about the upper pivot. The inherent problem is that these designs lose rate near full jounce due to the suspension mechanism components becoming generally planer. That is, all the suspension components fold down until they are lying relatively flat as the suspension components move at full jounce. This is due to the large vertical travel requirements of a snowmobile suspension.

The result of the suspension components becoming planar is that the load vector of the LFEs begins to point horizontally instead of vertically. This transfers load into the internal components of the suspension and does not react vertically to suspend the vehicle.

Also, as the suspension components become planar, the moment arm through which the suspension reacts increases at a faster rate than can be controlled by the shock/spring ratio, regardless of the type of linkage used to accelerate the shock/spring. With reference again to FIGS.

1 and 2, FIG. 1 shows a 2D representation of suspension 10 at full rebound. 2 shows suspension 10 at full jounce. The front and rear LFEs 22, 24 become generally planar and lay down and point nearly horizontally in FIG. The rear torque arms get “longer” measured from the upper pivot to lower pivot in the horizontal direction.

Even with a complicated linkage to help stroke the rear LFE 24, a progressive rate cannot be maintained due to the two reasons listed above. This is true for all conventional snowmobile rear suspension systems. Load at the slide rails and, more importantly, the bias between front and rear load is directly related to coupling, especially for a front load case. Consider the traditional suspension as illustrated in FIG. The architecture is such that the front spring/damper 22 acts between the front arm 14 and the slide rail 12, and both the torsion springs and rear damper 24 act between the rear arm 16 and the slide rail 12 near the front. Therefore, during a front load case, both springs and dampers 22, 24 have a large effect on load and rate.

The same is true of a rear load case. Attempting to tune the front LFE 22 will change the load/rate at the front and rear, and vice versa.

Also if the coupling were increased, the rail angle decreases through travel and the rate will increase. In order to tune the suspension rate, the front LFE 22, rear LFE 24 and torsion springs, and coupling angle all need to be adjusted. To improve this system: 1) Front coupling can be used primarily to control front load/rate, 2) Front preload is adjusted by a small LFE near the front of the rail (has a very small affect on rate), and 3) rear preload and rate is determined by the rear arm only. To achieve this with actual architecture, the main rear LFE needs to react only at the rear arm and with no other suspension components.

Therefore reacting the LFE on the chassis in the above discussion is important not only for progressive rate, but also for load bias. When these three conditions are true, rear coupling does not greatly influence rate.

This is realized because the front LFE is only used for preload so there is generally very little rate to “borrow” from the front of the rail during a rear load case. The state of snowmobile rear suspensions in the industry consists entirely of falling rate, or regressive suspension designs. Even though there is a large variety in the suspension architecture from one manufacture to another, commercially available designs yield an overall suspension stiffness that decreases as the suspension is compressed toward full jounce. Some architectures yield discontinuities that may locally spike the rate for a short time (such as an overload spring), but afterwards the rate continues to decrease.

Because most design effort is directed at optimizing a damper or spring motion ratio instead of analyzing the entire suspension system there are currently no progressive rate suspensions in the industry. This conventional suspension arrangement poses two problems. First, track tension through suspension travel relies on the relative placement of the suspension arms and wheels to the drive shaft.

Suspension mount locations are often determined not only by specific, desired suspension characteristics, but also on track tension packaging. Problems are encountered from both an over and under tensioning track standpoint. Second, the front arm placement is limited to remain outside the drive sprocket diameter due to interference with drive train components. This creates problems when attempting to change the weight transfer behavior of the rear suspension, which is dominated by front arm mount location. With this design, the improved suspension may eliminate the carrier wheel. This changes the load vector into the suspension from the track due to tractive forces between the track and ground.

In the illustrated embodiment, the load vector from the track is more horizontal which induces less pitch and weight transfer than a traditional suspension. To improve this, the front arm is moved significantly forward to facilitate weight transfer. This point can move forward incrementally until it encounters the drive wheel inscribed circle. At this point, it can only move coaxial with the drive sprocket. The illustrated embodiment of the present invention utilizes a coaxial front arm mount as discussed herein to facilitate weight transfer and pitch. At least one linear force element (LFE) may be positioned between the chassis and the rear linkage. The at least one LFE may be positioned outside of the envelope defined by the endless track.

The at least one LFE may be positioned above the tunnel. The chassis is comprised of framing attached to the exterior of the tunnel, and comprises a central bracket to which one end of the LFE is attached. The snowmobile may further comprise a drive shaft mounted transversely through the tunnel, where the front linkage is attached to the tunnel, at a position coaxial with the drive shaft. The chassis may further comprise load arms directing forces along a load path which are coincident with a rotational axis of the drive shaft. The snowmobile may further comprise at least one front linkage operatively connecting the slide rails to one of the chassis or the tunnel; and at least one rear linkage operatively connecting the slide rails to the chassis. The at least one LFE is positioned between the chassis and the rear linkage. The at least one LFE is positioned outside of the envelope defined by the endless track.

The at least one LFE is positioned above the tunnel. The chassis may be comprised of framing attached to the exterior of the tunnel, and may comprise a central bracket to which one end of the LFE is attached. The snowmobile may further comprise a drive shaft mounted transversely through the tunnel at a position coaxial with the rotational axis. The chassis may further comprise load arms directing forces along a load path which is coincident with the drive shaft rotational axis. The chassis may comprise tubes, attached to and straddling the tunnel, where the tubes extend substantially vertically. The tubes define a steering hoop. The chassis may comprise tubes, attached to and straddling the tunnel, the tubes extending substantially vertically.

The tubes define a steering hoop. The snowmobile may further comprise at least one front linkage operatively connecting the slide rails to the tunnel; and at least one rear linkage operatively connecting the slide rails to the chassis. The snowmobile may further comprise a drive shaft mounted transversely through the tunnel, wherein the front linkage is attached to the tunnel, at a position coaxial with the drive shaft. The chassis may further comprise load arms directing forces along a load path which are coincident with the drive shaft rotational axis. The at least one LFE may be positioned outside of the envelope defined by the endless track. The at least one LFE may be positioned above the tunnel.

BRIEF DESCRIPTION OF THE DRAWINGS•. With reference first to FIGS.

8 and 9, an illustrated embodiment of a suspension 30 of the present invention is shown. Suspension 30 includes a pair of slide rails 12, and a swing arm 31 having a first end coupled to slide rails 12 at location 33 and a second end 35 pivotably coupled to the chassis. A frame 41 ( FIG. 9) is configured to define a tunnel 40 which receives the track 39 therein.

In the illustrated embodiment, the main LFE 32 is located generally horizontally above the tunnel 40 with one end connected to the chassis at location 34 and the other end 36 connected to a bell crank 38 that redirects the load vertically. A second LFE 37 includes a first end 43 which is pivotably coupled to the slide rails 12. A second end 45 of LFE 37 is pivotably coupled to link 47.

As shown in FIG. 8, an opposite end of link 47 is pivotably coupled to swing arm 31 by connector 49. As discussed below bell crank 38 generally forms a triangular shape with corner 50 coupled to the chassis, corner 52 coupled to end 36 of LFE 32, and corner 54 coupled to end 45 of LFE 37. As discussed above, the main LFE 32 is illustratively placed outside the envelope defined by track 39 and above the tunnel 40 as shown in FIG. One end 34 of the LFE 32 is connected to the chassis and the other end 36 is connected to an end of a bell crank 38. The other end of the bell crank 38 is connected to the suspension components, either through a link, pivot, slider, or other suitable connection. The suspension components extend around the track 39 in order to connect components located within the envelope of the track 39 to the LFE 32.

Details of an illustrated embodiment of this connection are described below. As shown in FIG. 9, one such improved suspension places at least one LFE outside the envelope defined by the track and above the chassis tunnel.

One end of the LFE is connected to the chassis and the other end is connected to an end of a bell crank. The other end of the bell crank is connected to the suspension components, either through a link, pivot, slider, or other suitable connection. As the suspension compresses into jounce, the suspension end of the bell crank moves vertically some amount which causes the crank to rotate.

Rivers And Floodplains Ebook Reader on this page. This, in turn, causes the LFE end of the bell crank to move horizontally and stroke the LFE. This is what provides the vertical suspension rate.

Comparing an example of the improved suspension at full rebound ( FIG. 10) to full jounce ( FIG. 11), it is shown that the horizontal distance from the crank pivot to the crank suspension end increases. This increase of the “output” bell crank moment arm by itself would make the vertical load decrease through travel.

However, the vertical distance between the crank pivot and the crank LFE end increases. This increase in the “input” bell crank moment arm balances the increase in the “output” and maintains the vertical load. By changing the relative length of the arms and the angles between them, a progressive rate can be generated for most suspension load cases. The arm lengths and angles of the bell crank 38 are important to the operation of the suspension 30. 10 and 11 show the illustrated embodiment at full rebound and at full jounce, respectively.

The large triangle represents the bell crank 38. The left corner of the triangle 50 below the LFE 32 is the pivot connector to the chassis. The top most corner 52 is connected to the LFE 32, and the bottom most corner 54 is connected to the suspension (in this case through the link 47). Comparing the suspension 30 at full rebound ( FIG. 10) to full jounce ( FIG. 11), it is shown that the horizontal distance from the crank pivot to the crank suspension end increases.

This increase of the “output” bell crank moment arm by itself would make the vertical load decrease through travel. However, the vertical distance between the crank pivot and the crank LFE end increases. This increase in the “input” bell crank moment arm balances the increase in the “output” and maintains the vertical load. By changing the relative length of the arms and the angles between them, a progressive rate can be generated for most suspension load cases. As stated above, the present invention may also be applied to existing rear suspensions, and FIGS.

13 and 14 show a retro-fit to the Polaris Fusion® rear suspension shown in FIGS. 13 discloses use of a single bell crank 38 coupled between rear suspension arm 16 and LFE 32. 14 discloses use of dual bell cranks 38 and 38′.

14, a first end of bell crank 38 is coupled to end 36 of LFE 32 and a second end of bell crank 38 is coupled to rear suspension arm 16 by link 51. A second bell crank 38′ has a first end coupled to end 34 of LFE 32 and a second end coupled to front suspension arm 14 by link 51′. With reference first to FIG. 15-17, rear suspension 60 is shown attached to tunnel 40, and illustrates the suspension 60 coupled to a frame 41 which defines the tunnel 40 for track 39 ( FIGS. This system is generally comprised of, slide rails 12, LFE 32, bell crank 38, tunnel 40, chassis 70, front swing arms 80 and an A-shaped pivot member 96. More particularly, LFE 32 is shown suspended between bell crank 38 and a chassis structure 70. Bell crank 38 is attached to A-shaped pivot member 96, which in turn is attached to slide rails 12.

17 shows that the main LFE 32 is located horizontally above the frame 41 which defines the tunnel 40. In the embodiment of FIGS. 15-17, and as best shown in FIGS.

16 and 17, a front swing arm 80 is pivotably coupled to the slide rails 12 by a pivot connection 82 as discussed in further detail below. An opposite end 84 of swing arm 80 is pivotably coupled to the chassis about which a drive mechanism 85 is attached, having an axis which is coaxial with a drive shaft 86 as also discussed in detail below. With reference first to FIGS. 18 and 19, the connection of the front swing arm 80 to slide rail 12 will be described. As mentioned above, conventional suspension arrangements pose two problems.

First, track tension through suspension travel relies on the relative placement of the suspension arms and wheels to the drive shaft. Suspension mount locations are often determined not only by specific, desired suspension characteristics, but on track tension packaging. Problems are encountered from both an over and under tensioning track standpoint. Second, the front arm placement is limited to remain outside the drive sprocket diameter due to interference with drive train components. This creates problems when attempting to change the weight transfer behavior of the rear suspension, which is dominated by front arm mount location. There are two illustrated arrangements in which the arm 80 is mounted coaxial to the drive shaft 86 either on the drive shaft 86 or the chassis. 18 shows the first arrangement where the arm 80 is mounted directly to the drive shaft 86.

In this arrangement, bearings are used in the connection to allow the drive shaft 86 to rotate within the ends 84 of the suspension arm 80. The advantages of this connection are twofold, lateral packaging of the arm 80 in the chassis tunnel is easier, and the arm strengthens the drive shaft 86. In this embodiment, however, high speed bearings are required at this connection, and the drive shaft 86 must now react to suspension loads. Traditional snowmobiles have typically used drive shafts that are wider than the tunnel.

This is to simplify the number of parts in the assembly and still allow mounting to each edge of the tunnel with a single shaft. However, this makes assembly and service difficult. In order to remove the drive shaft you need to open the chain case, loosen the drive shaft bolt, slide the drive shaft out of the chain case, twist the drive shaft and remove it from the tunnel. Sliding the drive shaft and twisting to the side can be very difficult due to the tunnel/track clearance.

This first sleeve embodiment is depicted in FIGS. 20 and 21, and includes a drive shaft similar to current designs, but the drive sprockets 88 are mounted to an outer sleeve 180 (instead of the shaft directly) that is slightly narrower than the tunnel 40.

The two parts are then torsionally coupled through either sliding splines, hexes, or other similar fit. The inner shaft 182 is tightly mounted to the chain case 184 by means of a fastener and the outer sleeve 180 is compressed when the inner shaft 182 is tightened from the end opposite the chain case 184.

To assemble this design, and as best shown in FIG. 21, the sleeve 180 is placed in the tunnel and the shaft 182 slides completely through the sleeve 180, from the outside of the tunnel, into the chain case 184. The shaft 182 is torsionally coupled to the drive mechanism inside the chain case and fastened solidly with a screw 186. The chain case 184 has an access opening 187 to install the screw 186 so the case 184 does not need to be opened to access the drive shaft 182.

The entire assembly is clamped tight from the side opposite the chain case 184. As this is tightened, the outer sleeve 180 is compressed from each end by the main drive shaft bearings. To assemble this design, and with reference to FIG.

23, the drive shaft 188 is positioned slightly off center from the tunnel, enough for the chain case end of the shaft to clear the tunnel wall. A notch is present in the tunnel wall for the free end (non-chain case end) of the shaft to pass through into the correct position. The drive shaft 188 is then moved toward the chain case 184 and pilots on the case bearing.

The spline stub 190 is then inserted from the outside of the chain case 184 and torsionally couples the drive shaft 188 to the chain case drive mechanism 192. An access hole 187 is present in the case cover so the case does not need to be opened to install or remove the stub 190. A fastener 191 is then threadably received in the end of the shaft 188, closest to the case 184, clamping the drive shaft 188 to the chain case. This fastener 191 is then enclosed by a cover 194 for the access hole 187.

Lastly, the free end of the shaft 188 is tightened against the main drive shaft bearing. The suspension of the present invention packages the suspension around the track. That is, the track actually passes through one or more suspension components.

This design yields superb track tension values throughout travel. Due to a lack of a carrier (upper) track wheel, and coaxial mounting of the swing arm and drive sprocket, the tension in the illustrated embodiments only relies on the drive sprocket wheel 88 and idler wheels 108 to keep the track tight to prevent “unwrapping” around the rail bend profile as shown in FIGS.

Track tension is easily tuned by sizing the idler wheel 108 with the drive wheel 88. Therefore, elimination of carrier wheel in conjunction with coaxial swing arm mounting greatly simplifies track tensioning in the illustrated embodiments. With respect now to FIG. 24, A-shaped rear pivot 96 will be described in greater detail. As mentioned above, A-shaped rear pivot 96 connects bell crank 38 to LFE 32. A-shaped rear pivot 96 is shown pivotably coupled to bell crank 38 by connection 98. A first arm 100 of pivot 96 is pivotably coupled to slide rails 12 at location 102.

As shown best in FIG. 25, second arm 104 of pivot 96 is coupled to slide rails 12 by a coupling slider 106, having an arced slot 107 that facilitates coupling between the front and rear. A block 105 coupled to arm 104 moves back and forth in slot 107.

The following outlines the function of each component in the embodiment shown in FIGS. The swing arm 80 is pivotally connected to the chassis coaxial with the drive shaft 86, low and forward on the slide rail 12 to facilitate weight transfer. The pivot 96 is pivotally connected to the slide rail 12 near the rear, and to an arced slot 107 that facilitates coupling. The pivot 96 is “locked” to the slide rail 12 at the extents of the slot 107. The geometry is coupled to the front when the pivot is at the bottom, to the rear when the pivot is at the top.

The crank 38 is pivotally connected to the pivot 96 at one end 54 and the chassis at the other end 50. The crank 38 acts as the rear arm of the four-link. The preload spring 90 is connected between the swing arm 80 and the slide rail 12. This spring 90 is used for preload bias and does not appreciably affect rate. The main spring/damper 32 is connected between the crank 38 and the chassis.

The location on both determines how progressive the suspension is. As snowmobiles develop, accommodations in the chassis must be made for faster, more powerful engines, longer travel suspension, more precision handling, and improved durability. This means the chassis must be stronger and stiffer. The most intuitive method to increase strength and stiffness is to directly connect the suspension hard points with more significant structure than a thin walled tunnel can provide. The result is a direct load path between the front suspension mounts, the rider input points, and the rear suspension mount points, such that the load path can only terminate in a structurally durable member of the chassis. The chassis structure, especially in the rear section of the snowmobile, becomes considerably more important when the LFE reacts outside the suspension, as described in the above discussion.

In this case, rear suspension loads are not only internal to the suspension, but are directed into the chassis such that the chassis structure is an integral part of the suspension. As discussed above, a suspension system is described for support for the LFE 32 above the tunnel 40. The sub frame 70 was shown in FIG. 15 for mounting LFE 32 above tunnel 40. With reference now to FIG.

27, the snowmobile sub frame 70 will be described for mounting and supporting the LFE 32 above the tunnel 40. In the embodiment of FIG.

27, a steering hoop 130 is mounted to opposite sides of the frame 41 (see FIG. 9 and 15) by fasteners 132. Clevis brackets 134 are provided for coupling to opposite sides of the bell crank 38 ( FIG. A central bracket 136 is provided for coupling to end 34 of LFE 32 ( FIG. Four support arms 138 hold the bracket 136 in place as best shown in FIG. Support arms 135 extend from brackets 134 to brackets 137 on opposite sides of the tunnel frame 41 as shown in FIG. Each arm 135 directs forces along a load path from the bell crank 38 along a line 139 (shown in FIG.

15 and 28) which passes through the axis or rotation of the drive shaft 86. 28 illustrates that a plurality of load paths are directed through the axis of rotation of drive shaft 86.

A traditional snowmobile chassis relies solely on the tunnel frame assembly 41 to provide support for the rear suspension. Modern performance snowmobiles are reaching levels of performance at which a stiffer chassis would be ideal. By using a frame to attach directly to the pivot points of the suspension, and tie into existing structure found at the steering hoop 146, the support structure of the rear suspension is made much stiffer. The tunnel frame 41, while still partly supporting the rear suspension, is primarily acting as a track shield and foot support. An advantage of this structure is the direct load paths between the LFE mount, the rear arm mount, and the front arm mount.

Because the front arm is mounted coaxial with the driveshaft, the drivetrain (such as gearcase or transmission) also needs to be structural and becomes an integral part of the chassis structure. With this system, the tunnel itself may or may not be important to the overall chassis strength. If the tunnel was not structural, it would only acts as a snow shield and foot support.

The illustrated design features of the architecture of the rear suspension disclosed herein are summarized as follows: the main shock and damper (LFE 32) are mounted above the track 39 and above the tunnel 40 and react on the chassis. The chassis structure disclosed with reference to FIGS. 27-32 facilitate the over tunnel LFE design.

At least one suspension arm mounts to the chassis above the track, and the track passes through at least one suspension component. In other words, at least one component “wraps” around the track. The suspension 60 does not have a carrier wheel which yields a triangular track wrap path.

Sliders and bumpers are used to control the track direction, but these are not normally in contact with the track. Swing arm 80 mounts to the chassis coaxial with the drive shaft 86. Track 39 and drive shaft 86 are part of the suspension subsystem. They are installed and removed from the vehicle as one unit.

A slider slot 107 is used to control the relative angle of the pivot 96 to the slide rail 12 as shown in FIG. A block 105 coupled to arm 104 moves back and forth in slot 107. This controls both front-to-rear and rear-to-front coupling. More simple bumpers may be used on the slide rail instead of a slot 107, but the slot 107 offers lateral and longitudinal stiffness. A deflector shield 11 0 is mounted to idler wheel assembly 108 as shown in FIGS.

By mounting directly to this assembly only, the shield 110 moves with the idler wheel 108 when setting track tension. The vertical difference between the front and rear arm chassis mounts is illustratively 20% or more of the chassis link length (D). Finally, A/B ratio of links A and B is illustratively 1.6 to 2.0 or greater. Referenced by Citing Patent Filing date Publication date Applicant Title Feb 14, 2011 Jul 24, 2012 Polaris Industries Inc. Snowmobile and an oil container therefor * Aug 31, 2009 Nov 20, 2012 Soucy International Inc.

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Traction assembly for vehicle Nov 30, 2009 Mar 17, 2015 Polaris Industries Inc. Snowmobile and rear suspension for snowmobile May 30, 2014 Sep 29, 2015 Skinz Protective Gear Remotely adjustable suspension coupling * Apr 13, 2012 Mar 22, 2016 Bpg Recreational Inc. Suspension assembly for personal tracked vehicle Jun 16, 2015 Aug 30, 2016 Skinz Protective Gear Remotely adjustable degrees of freedom for suspension coupling Jan 10, 2014 Nov 29, 2016 Polaris Industries Inc. Engine having active exhaust valve position control system and method Jul 29, 2015 Jan 10, 2017 Polaris Industries Inc. Snowmobile Nov 22, 2013 Nov 7, 2017 Polaris Industries Inc. Snowmobile * Jan 17, 2007 Sep 27, 2007 Polaris Industries Inc. Snowmobile rear suspension * Jan 19, 2009 Oct 29, 2009 Robert Bessette Traction assembly for a vehicle * Nov 30, 2009 Mar 25, 2010 Polaris Industries Inc.

Snowmobile and rear suspension for snowmobile * Aug 31, 2009 Mar 3, 2011 Robert Bessette Displacement Limiting Assembly for a Track System * Feb 14, 2011 Jun 9, 2011 Polaris Industries Inc. Snowmobile * Mar 24, 2011 Sep 27, 2012 Robert Bessette Traction assembly for vehicle * Apr 13, 2012 May 22, 2014 Ryan James Fairhead Suspension assembly for personal tracked vehicle Dec 16, 2010 Aug 4, 2011 Polaris Industries Inc. Vehicle cooling system.

In finance, a foreign exchange option (commonly shortened to just FX option or currency option) is a derivative financial instrument that gives the right but not the obligation to exchange money denominated in one currency into another currency at a pre-agreed exchange rate on a specified date.[1] See Foreign exchange derivative. The foreign exchange options market is the deepest, largest and most liquid market for options of any kind. Most trading is over the counter (OTC) and is lightly regulated, but a fraction is traded on exchanges like the International Securities Exchange, Philadelphia Stock Exchange, or the Chicago Mercantile Exchange for options on futures contracts. The global market for exchange-traded currency options was notionally valued by the Bank for International Settlements at $158.3 trillion in 2005 For example, a GBPUSD contract could give the owner the right to sell?1,000,000 and buy $2,000,000 on December 31.

In this case the pre-agreed exchange rate, or strike price, is 2.0000 USD per GBP (or GBP/USD 2.00 as it is typically quoted) and the notional amounts (notionals) are?1,000,000 and $2,000,000. This type of contract is both a call on dollars and a put on sterling, and is typically called a GBPUSD put, as it is a put on the exchange rate; although it could equally be called a USDGBP call. If the rate is lower than 2.0000 on December 31 (say 1.9000), meaning that the dollar is stronger and the pound is weaker, then the option is exercised, allowing the owner to sell GBP at 2.0000 and immediately buy it back in the spot market at 1.9000, making a profit of (2.0000 GBPUSD? 1.9000 GBPUSD)? 1,000,000 GBP = 100,000 USD in the process. If instead they take the profit in GBP (by selling the USD on the spot market) this amounts to 100,000 / 1.9000 = 52,632 GBP. Although FX options are more widely used today than ever before, few multinationals act as if they truly understand when and why these instruments can add to shareholder value.

To the contrary, much of the time corporates seem to use FX options to paper over accounting problems, or to disguise the true cost of speculative positioning, or sometimes to solve internal control problems. The standard clich? About currency options affirms without elaboration their power to provide a company with upside potential while limiting the downside risk. Options are typically portrayed as a form of financial insurance, no less useful than property and casualty insurance. This glossy rationale masks the reality: if it is insurance then a currency option is akin to buying theft insurance to protect against flood risk. The truth is that the range of truly non-speculative uses for currency options, arising from the normal operations of a company, is quite small.

In reality currency options do provide excellent vehicles for corporates' speculative positioning in the guise of hedging. Corporates would go better if they didn't believe the disguise was real. Let's start with six of the most common myths about the benefits of FX options to the international corporation -- myths that damage shareholder values. Historically, the currency derivative pricing literature and the macroeconomics literature on FX determination have progressed separately.

In this Chapter I argue the joint study of these two strands of literature and give an overview of FX option pricing concepts and terminology crucial for this interdisciplinary study. I also explain the three sources of information about market expectations and perception of risk that can be extracted from FX option prices and review empirical methods for extracting option-implied densities of future exchange rates. As an illustration, I conclude the Chapter by investigating time series dynamics of option-implied measures of FX risk vis-a-vis market events and US government policy actions during the period January 2007 to December 2008. Chapter 2: This Chapter proposes using foreign exchange (FX) options with different strike prices and maturities to capture both FX expectations and risks.

We show that exchange rate movements, which are notoriously difficult to model empirically, are well-explained by the term structures of forward premia and options-based measures of FX expectations and risk. Although this finding is to be expected, expectations and risk have been largely ignored in empirical exchange rate modeling. Using daily options data for six major currency pairs, we first show that the cross section options-implied standard deviation, skewness and kurtosis consistently explain not only the conditional mean but also the entire conditional distribution of subsequent currency excess returns for horizons ranging from one week to twelve months. At June 30 and September 30, the value of the portfolio was?1,050,000. Note, however, that the notional amount of Ridgeway's hedging instrument was only?1,000,000.

Therefore, subsequent to the increase in the value of the pound (which is assumed to have occurred on June 30), a portion of Ridgeway's foreign currency exchange risk was not hedged. For the three-month period ending September 30, exchange rates caused the value of the portfolio to decline by $52,500. Of that amount, only $50,000 was offset by changes in the value of the currency put option. The difference between those amounts ($2,500) represents the exchange rate loss on the unhedged portion of the portfolio (i.e., the 'additional'?50,000 of fair value that arose through increased share prices after entering into the currency hedge). At June 30, the additional?50,000 of stock value had a U.S. Dollar fair value of $45,000.

At September 30, using the spot rate of 0.85:1, the fair value of this additional portion of the portfolio declined to $42,500. Ridge way will exclude from its assessment of hedge effectiveness the portion of the fair value of the put option attributable to time value. That is, Ridgeway will recognize changes in that portion of the put option's fair value in earnings but will not consider those changes to represent ineffectiveness. Aitan Goelman, the CFTC’s Director of Enforcement, stated: “The setting of a benchmark rate is not simply another opportunity for banks to earn a profit. Countless individuals and companies around the world rely on these rates to settle financial contracts, and this reliance is premised on faith in the fundamental integrity of these benchmarks. The market only works if people have confidence that the process of setting these benchmarks is fair, not corrupted by manipulation by some of the biggest banks in the world.” The Commission finalized rules to implement the Dodd-Frank Wall Street Reform and Consumer Protection Act regarding Regulation of Off-Exchange Retail Foreign Exchange Transactions and Intermediaries. The Commission also finalized Conforming Changes to existing Retail Foreign Exchange Regulations in response to the Dodd-Frank Act.

Additional information regarding these final rules is provided below, including rules, factsheets, and details of meetings held between CFTC Staff and outside parties.